Deep well sonic pumping process and apparatus



Feb. 14, 1967 A. G. BOBINE, JR 3,303,782

DEEP WELL SONIC PUMPING PROCESS AND APPARATUS Filed NOV. lO, 1965 United States Patent O 3,303,782 DEEP WELL SONIC PUMPING PROCESS AND APPARATUS Albert G. Bodine, Jr., Los Angeles, Calif. (7877 Woodley Ave., Van Nuys, Calif. 91406) Filed Nov. 10, 1965, Ser. No. 507,205 9 Claims. (Cl. 103-1) This invention relates generally to the eld of sonic deep well pumps and processes for operating such pumps, such as typically disclosed in my prior Patents Nos. 2,553- 541 and 2,702,559, and is particularly concerned with improvements to overcome certain problems encountered in practical field experience and which have proved to be especially resistant to solution, particularly under field environment conditions in which a number of pumps of this class have now been operated for a considerable period of time. These pumps as heretofore known have, in certain field conditions and environment, displayed a proneness to operational instability, and they have also often 'been found subject to lfatigue failure of the pump tubing string when used at high repetitive stress owing to poorly controlled sonic compressional wave action in the tubing string ofthe pump. The present invention provides `solutions to these problems which upon experimental tests have been found surprisingly effective.

Sonic pumps of the type to which the invention is applicable utilize sonic wave energy transmission down a long elastic line or column extending from 4ground surface to the bottom of the well, and which may be either the pump tubing string, a rod string inside the pump tubing, or a cable. Patent No. 2,702,559 disclosed the use of the pump tubing string as the elastic column which serves as the transmission line for the sonic energy, and the particular pump chosen for illustrative purposes herein is of that type, but without implied limitation thereto. lt is thus to be understood that the present invention is also `broadly applicable as well to sonic -pumps of the type in which a rod string or column inside the pump tubing serves as the elastic energy transmission line (Patent No. 2,553,541), as well as in case of substitution of an elastic cable for the rod string. In all these cases, the fluid-impelling action is imparted to the fluid in the tubing string by a series of check valves which open and close automatically in relation to sonic wave or vibration transmission occurring along the elastic column or vibration transmission line.

Also, in all cases, the sonic energy necessary for the operation of the pump is delivered to the elastic column, whether the column be the tubing string, a rod string, or a cable, from a mechanical oscillator or cyclic force generator which is of a mass reactance type and which is afforded with an acoustic coupling to said column. The simplest form of mass reactive oscillator is one employing a pair of rotating unbalanced weights, such as disclosed in my Patent No. 2,702,559. These unbalanced weights, turning on shafts with support bearings on the oscillator body, exert upon the oscillator body an alternating reactive or inertial force, which is then directly impressed upon the upper end of the sonic column. Such a mass reactive or inertial type force Igenerator or oscillator has the advanage that its output stroke accommodates itself automatically to the impedance of the sonic column driven thereby, and by operating at a frequency approaching resonance, force wastage in reciprocating certain masses of the system is counteracted by elastic compliance reactance of the sonic column, so that useful output from the oscillator is maximized. The alternating force output from the oscillator acts on the upper end of the sonic column to transmit down it successive waves of relative compression and tension. Preferably, as indicated above, these waves are generated by the oscillator at a resonant standing wave frequency of the sonic column, causing the check vales to undergo cyclic displacements longitudinally of the liquid column such as to impel the liquid column up the tubing, as all set forth in my aforesaid Patent No. 2,702,559, and ywhich is incorporated herein `for a disclosure of the basic sonic pumping process fby this reference.

Briefly, the usual sonic pump, and the type disclosed herein for illustrative purposes, utilizes the pump tubing string, which is composed of steel, and therefore highly elastic, as the sonic column, as aforesaid, and suspends this tubing string from a vertically vibratory patform supported on flexible springs. The oscillator, of the massreactive or inertial unbalanced-weight or rotor type, comprises a body or frame which supports the unbalanced weights, and which is mounted -directly onto the upper end of the tubing string. The unbalanced weights, by their rotation, create centrifugal forces whose horizontal components are balanced out and whose vertical components are additive. The resultant vertically oriented force or impulse is exterted through the bearings for the unbalanced wei-ghts or rotors and thus impressed on the oscillator body or housing; the oscillator body or housing then, in turn, exerts a cyclic inertial force on the upper end of the pump tubing, whereby the sonic wave action in the latter is established. This sonic wave action in the tubing string involves high accelerations at velocity maximum points spaced a half-wavelength apart therealong, and the check valves which function as iluid impellers, are mounted along the columns such that these accelerations are available for moving the check valves and irnpelling the iiuid -up the conduit of the tubing string.

The present invention is effective in a sonic pumping system such as described above to provide stability for the frequency and amplitude of the sonic wave transmitted along the column, to provide reliable automatic phasing for the opening and closing of the lluid-impelling valves, and to provide stable frequency monitoring and control of the mass reactive oscillator. Accordingly, the attainment of these benefits are among the general objects of the invention.

Factors concerning stability of resonant wave transmission in a column were iirst investi-gated experimentally in connection with my studies of sonic earth bore drills and sonic oscillators intended for industrial material processing. I have found however, as a result of subsequent experiments with sonic pumps in the actual eld environment of deep oil wells, that a number of special advantages accrue from stabilizing of the sonic wave pattern along the sonic elastic column of a deep Well sonic pump, and also limitation of the amplitudes of the stress and strain cycles in the elastic column of the pump.

In considering a sonic deep well pump of the character here under consideration, the underlying characteristic that most basically controls its performance is the feature of resonance concerned in the sonic wave pattern established along the elastic column. One outstanding property of a sonic pump operated at resonance is that, assuming use of a mass reactive oscillator for the drive of the system, operation in the region of resonance maximizes the amplitude of the elastically vibratory motion produced in the sonic column, i.e. the cyclic strain in the column, and therefore maximizes the fluid-impelling action of the check valves. It also leads to maximized cyclic stress amplitude in the sonic column, and sometimes to strain magnitudes which cannot be tolerated, and which are limited by the present invention as will later appear. Under a given set of conditions, peak resonance, and therefore the maximum of vibration amplitude, or

cyclic strain, is attained at some particular frequency. The magnitude of this vibration amplitude or strain drops off smoothly but very sharply as the frequency departs either above or below the frequency for peak or maximum amplitude. That is to say, starting with an oscillator frequency below that for peak resonance, and therefore peak cyclic strain amplitude, it will be noted that amplitude of column `vibration or strain increases with frequency, until a maximum is attained at a peaking frequency. Then, as the frequency is increased still further, the amplitude of vibration falls off again.

Also, because in deep wells the elastic column is quite long, the condition of resonance can occur at a fundamental frequency, and will repeat itself up the frequency scale, with frequency humps lat each harmonic frequency. At each such hump, there is encountered the above described sensitivity of amplitude as against frequency.

The present invention is grounded upon the discovery that with such a sonic pump, employing a mass reactive oscillator, the pumping action is especially stable if the drive frequency is held slightly on the low side of one of these resonant humps, either fundamental or overtone. The pumping action, and stroke amplitude of the vibratory column, proceed with stability and reliability if the frequency of operation is just slightly less than that of the closest resonance peak In other words, I have found that this type of sonic pump operates with unexpected resonant stability and reliability at a frequency such that any increase in frequency causes a Very appreciable increase in amplitude, and where a very substantial increase in frequency would take the operation of the pump over the point of peak amplitude at the next higher frequency for resonance. It is accordingly an object of the invention to achieve a stability factor in this way.

Consideration of the operation of the pump when managed so that the frequency of sustained operation is held just under that for a resonance peak leads to the conclusion that the opening and closing actions of the automatic valves yare materially benefited when made subject to this novel control. It should be noted that these valves should be closed during the phase of their cycle when they are applying pump-impelling energy to the fluid. Moreover, these valves must be open for free ow during the phase of the cycle when they are moving in opposite direction to the flow. In other words, in such a sonic pump the uid is impelled by the elastic column causing the valve to move, While closed, in the direction which impels the fluid. This is when the elastic motion of the column is in the direction of fluid ow. On the return phase of this elastic motion cycle, when the valve is moving in opposition to tluid ow, the valve should be open, so `as not to impede the fluid flow. This desirable phasing of the valves is effected automatically to a very high degree of continuously stable performance :by establishing, through controlled drive of the oscillator, of elastic wave vibration in the column immediately on the low side of a resonance peak, as above described.

It is accordingly `an object and an accomplishment of the invention to obtain valve pumping effectiveness and efficiency as well as resonant stability by operating on a low side of a resonant frequency.

As mentioned hereinabove, the oscillator used in connection with the sonic pump is a type employing a cyclically movable mass, and may be characterized as of a mass reactive or inertial type. The inertia reactance resulting from confinement of the cyclically movable mass to its orbital path creates a cyclic force of inertia within the frame or body of the oscillator, and this force is applied to the above described elastic column. The coupling between the two is a mass-reactive type of acoustic coupling, with characteristics as mentioned hereinabove. It is an energy coupling whereby the output energy of the prime mover which drives the mass-reactive oscillator is coupled in energy transmission relationship to the resonantly vibratory elastic column. I have found that a steady reliable frequency of operation for such a massreactive type of oscillator can be established and maintained by using an operating frequency which is on the low side of a resonance peak. It is accordingly an object and an accomplishment of the invention to improve frequency stability for the oscillator by operating on the low side of resonance.

Employing the practice of the present invention, I can also easily adjust the actual amplitude of elastic vibration in the elastic column by adjusting t-he prime mover to produce a frequency in the low side resonance range, i.e. in the region of resonance, but below the frequency for the resonance peak, where the desired stroke or strain amplitude is attained. Under these conditions the amplitude tends to stay desirably stable and within reasonable limits of stress application to the sonic column. This purpose and accomplishment is an important object and feature of the invention.

An especially significant feature and advantage arising out of the practice of establishing operation in the low side resonance range is that, in the event of a sudden decrease in pumping load, s-uch as a large gas bubble going through some valves which have been working against a fairly rm liquid column load, the system will then not tend to overspeed and jump the resonant peak. This is because by going up to a higher frequency, as the result of loss of load, the prime -mover experiences regain of load beca-use the amplitude increases with frequency, when starting on the low side. This increase in amplitude results in picking up and impelling more uid load, even with some gas in the liquid, and the oscillator and its prime mover are thus still held within a close range of frequency or speed, and still on the low side of the resonant peak. In other words, with the practice of the invention, the system does not tend to run away in speed, and damage itself, even though the long elastic column, vibrating in the resonant range, tends, runder changing load conditions, to exert an influence on the prime mover as regards frequency or speed. Thus, under the conditions of the invention, the resonantly vibratory col-umn is prevented from dominating the prime lmover, by reason of the above noted increase in amplitude and therefore pick-up of additional load whenever there has been a drop in load. Operation on the low side of resonance thus introduces an 'automatic stabilizing characteristic which is not present at the peak of resonance, nor on the downslope above the peak of resonance.

A problem which I have encountered in eld experience with sonic pumps using elastic vibratory pump tubing is that, wit-h large repetitive cyclic stresses set up in the pump tubing when vibrating in a resonant standing wave pattern, `under conditions permitting frequency rise to the very peak of resonance, the tubing string can be damaged, and sometimes parts, with ensuing complete shut-down of the well until a time-consuming repair can be accomplished. In this connection, it should be understood that, first of all, commercially available pump tubing is not designed, nor composed of the necessary highquality alloy steel, to withstand the repetitive stresses likely to be encountered under peak resonance conditions with an adequate factor of safety. In practice, `under certain environmental conditions, or practical design or setup of equipment, complete tubing failure is sometimes experienced.

Tubing can actually fail in either of two ways under high stress. First, under a longitudinal stress f/A, where f is the applied longitudinal force and A is the crosssectional area of the tubing, there will be a longitudinal elastic strain equal to AZ YT where Y is Youngs modulus, l is the length of a halfwavelength segment of the tubing string when not undergoing vibration, and Al is the change in length in such segment resulting from the imposed stress. If the stress f/A is sufficiently high, the elastic limit is exceeded, and the yield point reached, at which the tubing does not make a total elastic return to its original length when the stress is removed. It is clear that strain to or beyond the elastic limit cannot be tolerated, and it is thus important to prevent the attainment of `peak resonance such as can lead to such amplification of stress and strain to or beyond the elastic limit.

Additionally, aud of an even more serious nature, is the problem of an occasional fatigue failure of the tubing string owing to la repetitive stress beyond the endurance limit of the tubing. The endurance limit in the present instance may be defined as the maximum repetitive stress amplitude which the ytubing string can withstand indefinitely without fatigue failure. It has been my experience that the best commercially available pump tubing has, at best, a factor of safety against such fatigue failure that must be considered as inadequate when operation at peak resonance frequency is tolerated, using an oscillator otherwise of no greater output force than is required to assure delivery of sufficient sonic energy to the tubing and 'all the Way down to the lowerrnost valve therein. It has further been my experience that, in practice, the endurance limit of the tubing becomes a function of the mechanical condition of the tubing, particularly as regards notches and scratches in the surface of the tubing, as well as of the quality of the steel of which the tubing is composed. Moreover, any corrosive yaction of well fluids tends to lower the endurance limit of the tubing as time passes. The endurance limit of a particular tubing string under particular environmental conditions can easily be determined after some field experience, and, when known, the practice of the present invention is then directed to the limiting of the maximum stress and strain amplitudes in the tubing to 4a level consistent with the determined endurance limit. The invention thus contemplates limitation of the frequency of operation to a level sufficiently below that for peak resonance that the stress exerted in the tubing or other elastic column, and the resulting strain therein, will be no greater than that for the ascertained, or in some cases estimated, endurance limit of the tubing or column in the environment of the well. This of course contemplates operation in the realm of resonance, i.e. fairly well up the low side of the resonant hump, but safely below the peaking frequency where stresses and strains become resonantiy amplified to values hazardous to the pump tubing or other elastic wave transmission line. It is accordingly an object 'and a feature of the invention to hold the cyclic stress and strain of the elastic cycle to a value which is within the elastic endurance limit of the material of the tubing string or other elastic column, as affected by environmental conditions, often as determined by field tests. It is a further object to limit such stress and strain, and thereby avoid fatigue failure, by operating the pump at a frequency which is in the range of resonance, but below the frequency for peak resonance, and thus below maximized cyclic stress amplitude.

I have also discovered that the high cyclic stress amplitude has an eect upon the wave transmisison efficiency of the elastic tubing string or other column. Thus, if the elastic stress value is taken to a fairly high cyclic amplitude, as above 20,000 p.s.i. fiber stress for a steel material, then in such a long transmission line system the internal damping, or elastic hysteresis, in the material of the steel itself becomes very large. This condition results in a considerable attenuation of the sonic Wave in the deep well pumping system, and reduces pumping efficiency accordingly, I have discovered that a stress amplitude of 20,000 p.s.i. is a preferred upper limit, which I establish by correspondingly limit- 6 ing the strain amplitude or elastic stroke at 'a given frequency.

It is accordingly an object of the invention to keep resonant performance sufficiently on the low side of the frequency for peak resonant amplitude as to hold cyclic stress amplitude (plus and minus cyclic excursions from mean static stress) within the limit of substantially 20,000 ps1.

In sonic pumps, the elastic stroke or vibration 'amplitude in the elastic sonic Wave transmission line or column of course determines the cyclic motion of the fluid-impelling element. Therefore, because of the limiting of the stress amplitude in the column, as described above, the action of the fiuid-impelling element is sometimes limited to a less than desirable amplitude for good pumping efiiectiveness. Actually, the problem arises because the cyclic pumping inliuence applied to the liquid column by each impeller has been curtailed by limitation of strain amplitude. However, it should be noted that this pumping impulse is also influenced by the rigidity of the liquid. That is to say, for a given cyclic stroke amplitude, or impelling element velocity amplitude, the more rigid the liquid` in the pump column, the greater will -be the kinetic energy, or pumping effort,` applied thereto by the impelling element. In other words, the sonic pump benefits from the use of a high-impedance liquid column in the pumping tubing. To explain further, the impedance Z of the liquid in this sense is equal to pc, where p is the density of the liquid and c is the velocity of sound in the liquid. I have found that a degassing of the liquid of the pumped liquid column, having the effect of increasing the density thereof, thereby improves its impedance and rigidity for improved pumping effect. The invention contemplates a degassing of the liquid before it enters the liquid column, such as by having an initial downfiow path of the liquid over the outside of the vibrating tubing before entry, or by degassinv a region of the liquid column within the pump, such as by having a perforation or vent in the tubing wall for escape of gas.

It is accordingly one object of the invention to irnprove the effectiveness of the sonic pumping action of a sonic pump by providing a more rigid liquid column.

There is, however, a limit to which the acoustic irnpedance of the liquid column can be raised by de-gassing, and I have found that the total kinetic energy ladded to the liquid Vby the sonic column can be infiuenced by the spacing distance of the impelling elements. The length of liquid column above a given impe-ller element which is accelerated by that impeller element is a function of liquid impedance. The lower the impedance of the liquid, the shorter the effective length of .liquid column accelerated by the sonically actuated impeller. I have discovered that there is no benefit, and actually a disadvantage by way of increased flow resistance, in having the impellers spaced closer than the length of column which can be accelerated by an impeller, taking into account the rigidity of the liquid column. Closer spacing simply means that some of the impellers are merely moving along with the liquid accelerated by the next impeller below,- and these impellers add flow resistance without any counteracting benefit. For most well liquids, including relative spongy crude oil, I prefer to have the impeller spacing no less than an average of five feet, and such spacing limitation is a preferred feature of the invention.

On the other hand, with some fairly high impedance liquids, eg., non-gassy petroleum wells, or water wells, the impeller spacing along the column can be considerably greater. However, if this impeller spacing becomes too great, even for the higher impedance liquids, there will be an intervening length of liquid column little affected by the impellers. This portion of the liquid then has to be pushed along .'by the elastic portion of the liquid column acted upon by the impeller, thus requiring more active impeller motion and correspondingly greater cyclic stress in the transmission line or column. I have found that performance tends to be undesirable with many well liquids if the spacing of the impeller is much more than one hundred feet, and accordingiy it is an object of the invention to provide for the average spacing of the fluid impellers, along an effective portion of the sonic column, to be no more than substantially one hundred feet.

Referring again to the matter of the acoustic impedance of the liquid, I have found that for some well fluids, going through the change in pressure environments `as they rise in the pump tubing string, the average effective impedance along the length of the column is sufficiently low that pumping is more effective if the actual cyclic stroke of the elastic vibration cycle in the sonic energy transmission line is fairly large. Since the cyclic stress and strain in the elastic column is a function of stroke or amplitude times frequency, the frequency must be relatively low in order for this stroke to be relatively large. Put in other language, if a long stroke is desired, this can be obtained easier and with less stress by use of a low operating frequency and hence a correspondingly long wavelength along the tubing. In such case, of course, i.e. with a long wavelength distance along the pipe, a given elastic stroke amplitude is attained with relatively low stress. I have found, in this connection, that there is a surprising advantage in practicing the sonic pumping process with the frequency under one hundred cycles per second, and it is accordingly one purpose of the invention to obtain a large stroke amplitude with relatively low stress in the tubing by having the frequency of the sonic wave in the tubing no greater than one hundred c.p.s.

On the other hand, the impeller applies energy to the liquid column by cyclically yaccelerating the liquid column. Since such acceleration, for a given stress in the transmission column, is a function of frequency squared, i.e. acceleration=(21rf)2x, where x represents cyclic stroke amplitude, the frequency must not be too low for good pumping effectiveness, and I have found that, for many well liquids and materials in the sonic wave transmission column, this frequency should not be less than substantially four cycles per second. It is accordingly a further object of the invention to assure adequate accelerating force by having the frequency at least as great as four c.p.s.

I have found that the attainment of a frequency of greater than four cycles per second can best be accomplished at most well depths by operating the oscillator at an overtone or harmonic of the fundamental resonant standing wave frequency of the sonic column or transmission line. It is accordingly an object to practice the invention using resonant overtone frequency in the elastic column.

Referring further to the above mentioned problem of establishing good frequency stability for the resonant standing wave pattern in the elastic transmission line or column, I have found a striking advantage by the employment, for the drive of the rotating unbalanced weight oscillator, of a drive shaft, with universal joints to permit bending displacement of the shaft axis, between the oscillator and its prime mover. Quite evidently there is a direct feedback effect from the sonic circuit composed of the resonant elastic column owing to the reactance of this column that causes the rotating unbalanced weights to tend to turn with other than constant angular velocity. Thus there is apparently a certain amount of non-linearity in the wave pattern of the column, land this can be caused, at least in part, because the impellers, being unidirectional in their application of energy to the liquid column, function as rectifiers in the equivalent circuit. In addition, since the oscillator is not operated at peak resonance, as before described, there is a phase angle between the sonic wave source, i.e. the oscillator, and

its acoustic load, the resonant elastic column. It accordingly follows that the oscillator tends to run in a non-linear manner, with alternating angular acceleration and deceleration, or in other words, without constant angular velocity. This of course contributes to instability of the sonic wave pattern in the elastic column.

I have found that the usual and so-called soft direct drive to the oscillator, such as a pulley and direct belt combination, permits the oscillator under these influences to meander, so to speak, through its cycle. On the other hand, I have found that a firm drive, such as a drive shaft, can be arranged to beneficially dominate the oscillator and hold it to a quite steady constant angular velocity. As mentioned, this can be accomplished by using a drive shaft to the oscillator, preferably one, of course, of large torsional stiffness. Of course, universal joints are needed in the drive shaft in order to permit freedom of vertical vibration of the oscillator in consonance with the vibration of the elastic column, while the prime mover, or its jack shaft, do not so oscillate. The torsional stiffness of this drive shaft assembly then holds the oscillator to relatively constant angular velocity with consequently more linear sonic wave output therefrom, and a more stabilized and effective wave pattern in the elastic column. Apparently, in this manner, the desired linear resonant sound wave pattern in the sonic column is hardened, so to speak, or in other language, strongly constrained to maintain itself against deviation in phase or amplitude, or wave form, such as otherwise appear and detract from the effectiveness by which the fluid-impelling elements deliver their kinetic energy to the liquid column. The use of such a drive shaft, in place of the more obvious belt and pulley drive, involves considerable additional expense, and is therefore preferably chosen over the simple belt and pulley system for the benefits mentioned, notwithstanding its added expense.

I have further found that the stabilizing effect described in the preceding paragraph can be further enhanced by use of a flywheel on the rotary drive system. A flywheel can of course be mounted directly on the oscillator; however, this undesirably adds mass to the vibrating oscillator, and accordingly, I find it lan advantage to install the flywheel on the drive shaft system rahead of the first or input universal joint. This drive shaft system can then be driven directly from the prime mover, or from the prime mover through a belt and pulley system to the drive shaft portion that carries the flywheel, for example.

Sonic discussion To aid in a full understanding of the phenomena of the present invention by those skilled in the acoustics art, and by others, the following general discussion, including definition of terms, is deemed to be of importance.

By the expression sonic vibration I mean elastic vibrations, i.e. cyclic elastic deformations, such as longitudinal, lateral, gyratory, torsional, etc., generated in a structure, or which travel through a medium with a characteristic velocity of propagation. If these vibrations travel longitudinally, or create a longitudinal wave pattern in a medium or structure having uniformly distributed constants of elasticity and mass, this is sound wave transmission. Regardless of the vibratory frequency of such sound wave transmission, the same mathematical formulae apply, and the science is called Sonics. In addition, there can be elastically vibratory systems wherein the essential features of mass appear as a localized infiuence or parameter, known as a lumped constant, and another such lumped constant can be -a localized or concentrated elastically deformable element, affording a local effect referred to variously as elasticity, modulus, modulus of elasticity, stiffness, stiffness modulus, or compliance, which is the reciprocal of the stiffness modulus. Fortunately, these constants, when functioning in an elastically vibratory system such as mine, have cooperating and mutual influencing effects like equivalent factors in alternating-current electrical systems. ln fact, in both distributed and lumped constant systems, mass is mathematically equivalent to inductance (a coil); elastic compliance is mathematically equivalent to capacitance (a condenser); and friction or other pure energy dissipation is mathematically equivalent to resistance (a resister).

Because of these equivalents, my elastic vibratory systems with their mass and stiffness and energy consumption, and their sonic energy transmission properties, can be viewed as equivalent electrical circuits, where the functions can be expressed, considered, changed and quantitatively analyzed by using well proven electrical formulae.

There have been other proposals involving exclusively simpie `bodily vibration of some part. However, these latter do not result in the benefits of my sonic or elastically vibratory action.

Since sonic or elastic vibration results in the mass and elastic compliance elements of the system taking on these special properties akin to the parameters of inductance and capacitance in alternating current phenomena, wholly new performances can be made to take place in the mechanical arts. The concept of acoustic impedance becomes of paramount importance in understanding performances. Here impedance is the ratio of cyclic force or pressure acting in the media to resulting cyclic velocity or motion, just like the ratio of voltage to current. In this sonic adaptation impedance is also equal to media density times the speed of propagation of the elastic vibration.

in this invention impedance is important to the accomplishment of desired ends, such as where there is an interface, or a coupling between two structures. A sonic vibration transmitted across an interface between two media or two structures can experience some reflection, depending upon differences of impedance. Impedance is also important to consider if optimized energization of a system is desired. If the impedances are adjusted to be matched somewhat, energy transmission is made very effective.

An important feature of these sonic circuits is the fact that they can be made very active, so as to handle substantial power, by providing a high Q factor. Here this factor Q is the ratio of energy stored to energy dissipated per cycle. In other words, with -a high Q factor, the sonic system can store a high level of sonic energy, to which a constant input of energy is respectively added and subtracted. Circuit-wise, this Q factor is numerically the ratio of inductive reactance to resistance. Moreover, a high Q system is dynamically active, giving considerable cyclic motion where such motion is needed.

Certain definitions should now be given:

impedance, in an elastically vibratory system, is, mathematically, the complex quotient of applied alternating force and liner velocity. It is analogous to electrical impedance. The concise mathematical expression for this impedance is where M is vibratory mass, C is elastic compliance (the reciprocal of stiffness, or of modulus of elasticity) and f is the vibration frequency.

Resistance is the real part R of the impedance, and represents energy dissipation, as by friction.

Reactance is the imaginary part of the impedance, and is the difference of mass reactance and compliance reactance.

Mass reactance is the positive imaginary part of the impedance, given by ZafM. It is analogous to electrical inductive reactance, just as mass is analogous to inductance.

Elastic compliance reactance is the negative imaginary part of impedance, given by Elastic compliance reactance is analogous to electrical capacitative reactance, just as compliance is analogous to capacitance.

Resonance in the vibratory circuit is obtained at the operating frequency at which the reactance (the algebraic sum of mass and compliance reactances) becomes zero. Vibration amplitude is limited under this condition to resistance alone, and is maximized. The inertia of the mass elements necessary to be vibrated does not under/ this condition consume `any of the driving force.

A valuable feature of my sonic circuit is the provision of enough eXtra elastic compliance reactance so that the mass or inertia of various necessary :bodies in the system does not cause the system to depart so far from resonance that a large proportion of the driving force is consumed and wasted in vibrating this mass. For example, a mechanical oscillator or vibration generator of the type normally used in my inventions always has a body, or carrying structure, for containing t-he cyclic force generating means. This supporting structure, even when minimal, still has mass, or inertia. This inertia could be a force-wasting detriment, acting as a blocking impedance using -up part of the periodic force output just to accelerate `and decelerate this supporting structure. However, by use of elasticaliy vibratory structure in the system, the eiect of this mass, or the mass reactance resulting therefrom, is counteracted at the frequency for resonance; and when a resonant acoustic circuit is thus used, with adequate capacitance (elastic compliance reactance), these `blocking impedances are tuned out of existence, at resonance, and the periodic force generating means can thus deliver its full impulse to the work, which is the resistive component of the impedance.

Sometimes it is especially `beneficial to couple the sonic oscillator at a low-impedance (high-velocity vibration) region, for optimum power input, and then have high impedance (high-force Vibration) at the work point. The sonic circuit is then functioning addition-ally `as a transformer, or acoustic lever, to optimize the effectiveness of both the oscillator region and the Work delivering region.

My invention employs, as the source of sonic power, a sonic resonant system comprising an elastic member in combination with an orbiting mass oscillator or Vibration generator, as above mentioned. This combination has many unique and desirable features. For example, this orbiting mass oscillator has the ability to adjust its input power and phase to the resonant system so as to accommodate changes in the work load, including changes in either or both the reactive impedance and the resistive impedance. This is a very desirable feature in that the oscillator hangs on to the load even as the load changes.

It is important to note that this unique advantage of the orbiting mass oscillator accrues from the combination thereof with the acoustic resonant circuit, so as to comprise a complete acoustic system. In other words, the orbiting mass oscillator is matched up to the vresonant part of its system, and the combined system is matched up to the acoustic load, or the job to be accomplished.

The combined system has a unique performance which is exhibited in the form of a greater effectiveness and particularly ygreater persistence in a sustained `sonic action as the work process proceeds or -goes through phases and changes of conditions. The orbiting 'mass oscillator, in this matched-up arrangement, is able to hang on to the load and continue to develop power as the sonic energy absorbing environment ch-anges with the variations in sonic energy absorption by the load. The orbiting mass oscillator automatically changes its phase angle, and therefore its power factor, with these chan-ges in the resistive impedance of the load.

A further important characteristic which tends to make the orbiting mass oscillator hang on to the load and continue the development of effective power, is that it also accommodates for changes in the reactive impedance of the acoustic environment while the work process continues. For example, if the load tends to add either inductance or capacitance to the sonic system, then the orbiting mass oscillator will accommodate accordingly. Very often this is accommodated by an automatic shift in frequency of operation of the orbiting mass oscillator by virtue of an automatic feedback of torque to the energy source which drives the orbiting mass oscillator. In other words, if the reactive impedance of the load changes this yautomatically causes a shift in the resonant response of the resonant circuit portion of the complete sonic system. This in turn causes a shift in the frequency of the orbiting mass oscillator for a given torque load provided by the power source which drives the orbiting mass oscillator.

All of the above mentioned characteristics of the orbiting mass oscillator are provided to a unique degree by this oscillator in combination with the resonant circuit. As explained elsewhere in this discussion the kinds of acoustic environment presented to the sonic source by this invention are uniquely accommodated by the combination of the orbiting mass oscillator and the resonant system. As will be noted, this invention involves the application of sonic power which brings forth some s-pecial problems unique to this invention, which problems are primarily a mat-ter of delivering effective sonic energy to the particular work process involved in this invention. The work process, as explained elsewhere herein, presents a special combination of resistive and reactive impedances. These circuit values must be properly met in order that the invention be practiced effectively.

The invention will be further described in connection with the following detailed description of a present illustrative embodiment of a sonic pump incorporating the invention and with which the invention may be practiced, reference for this purpose being had to the accompanying drawings, in which:

FIG. 1 shows in section, with longitudinal portions broken away, the underground installation of a sonic pump illustrating the practice of the invention;

FIG. 2 is a view partly in side eleva-tion and partly in longitudinal section, showing the above-ground equipment for the sonic pump of FIG. 1;

FIG. 3 is a plan view of the vibration generator or oscillator of FIG. 2;

FIG. 4 is a side elevational view of the oscillator of FIG. 3, being taken in accordance with the arrows 4-4 of FIG. 3;

FIG. 5 is a longitudinal section through a pump tubing coupling aud showing a check valve and fluid-impelling unit installed therein; and

FIG. 6 is a diagram showing a typical resonance curve, and showing also the range of operation characterized by the invention.

In the drawings an illustrative sonic oil well pumping installa-tion in accordance with the invention is shown, the well bore being shown to be lined by casing 10, surrounded by a cement annulus 10a, while annularly spaced inside casing 10 is the sonic wave transmission line and production tubing means, in this instance comprised of an elastic steel pump tubing string 12. The casing 10 is perforated as indicated at 13, and these perforations are shown to extend through the cement to the surrounding productive earth formation, all as clearly illustrated. The casing 10 extends downwardly into the well bore through a concrete slab 14 and a base plate 15, and at the top of casing 10, above ground level, is a casing head 16. The pump tubing 12 extends upwardly through the casing head and has mounted on its upper end a flow head 13, which in turn firmly mounts housing 29 of the vibration generator or oscillator designated generally at G. Flow head 18 has an outlet to which is coupled flow line 19, as illustrated. The oscillator G comprises gear housing 20 mounting certain unbalanced rotating weights which generate in said housing a resultant vertically oriented alternating or cyclic inertial force which is applied by the generator housing through the flow head 18 to the upper extremity of the pump tubing string 12.

The generator G as here shown comprises gear housing 20 affording bearings for two horizontal parallel shafts 23 and 24, which carry, inside said housing 20, meshing spur gears 25 and 26. Shafts 23 and 24 project from opposite sides of housing 20, and carry eccentric or unbalanced weights or masses 27, and of predetermined mass, and of eccentricity of center of mass, to afford the desired limited periodic elastic stress and strain amplitudes in the elastic column or transmission line (in this case, the elastic pump tubing string 12), when the oscillator is being driven at the predetermined frequency below a selected peak resonant frequency of the transmission line.

As here shown, the shaft 24 has coupled thereto, through universal joint 28, a drive shaft 29, and the latter is driven through universal joint 30 from a second drive shaft or jack shaft 32, which may be directly driven from the output shaft of a prime mover, but is here shown as driven from a prime mover in the form of an electric motor 36 through suitable pulleys and belts 37. The motor is shown mounted on a concrete base 38, which also supports frames 39 carrying bearings for the shaft 32. Shaft 32 preferably carries a flywheel 40.

AS can be seen in the drawings, the unbalanced rotary weights 27 are so phased as to move vertically in unison, so that the vertical components of the centrifugal forces generated thereby are additive. Also, the weights on the two shafts rotate in opposite directions, and move hori zontally equally and oppositely, each to its opposite number. Horizontal components of force are thus equalized and cancelled out.

Mounted on top of casing head 16 is a exible spring support device 41 for the tubing string and the oscillator G. This device 41 comprises a lower platform 42, which is mounted on a base flange 44 having a tubular neck portion 45 threaded within the casing head 16, and thus supported thereby. Coil compression springs 46 supported on platform 42 support, at their upper ends, an upper platform 47, and bolts 43 connect the two platforms 42 and 47 and limit their separation, but in an arrangement permitting vibratory movement of the upper platform relative to the lower one. The flow head 18 lis supported on upper platform 47, and oscillator G, xed to and supported by said flow head, and the pump tubing string 12 suspended therefrom, are thus spring supported by the springs 46 from the casing head and ground-supported casing. The springs 46 are relatively exible, and effectively isolate the casing from the vibratory movement of the oscillator and tubing string. In the operation of the pumping system, it is desired to establish a longitudinal resonant standing wave along the tubing string; and, to deliver good power into the transmission line, the transmission line should have a large stroke amplitude at its upper extremity. This means that a velocity antinode of the standing wave should be located at the upper end of the extremity of the tubing string where the oscillator is coupled, with a node one quarter-wavelength down from the upper end. With the large lumped mass of the oscillator at the upper end of the tubing string, however, the node rises toward the upper end, and a large stroke amplitude is not available. Accordingly, it is highly desirable, as taught in my Patent No. 2,902,937, to critically tune the vibratory system consisting of the oscillator G, the upper platform 47, and the springs 46, to resonate at the operating frequency of the oscillator G.

The neck 45 of base flange 44 is annularly spaced outside the pump tubing, and a threaded port 50 is preferably formed in said neck to provide communication with well annulus 52, for a purpose mentioned later.

The annulus 52 is sealed by tubular annulus seal 54 around the tubing above upper platform 42, being screwed tightly into base flange 44, and afforded at the top with a stuffing box or sealing means to the vibratory tubing means as indicated somewhat diagrammatically at 56.

Also, preferably, the spring mounting device 41 is equipped with a vibratory stroke or strain indicator, designated at 6i), and which comprises an arm 61 mounted on lower platform 42 and reaching to the vicinity of upper platform 47, where it is provided with a pointer 62. The amplitude of the vibratory stroke, i.e. of the vibrating elastic strain, can be ascertained by noting the stroke o'f the platform 47 in relation to this pointer. More accurate means for this purpose can be provided, but form no part of the present invention.

The tubing string comprises usual joints or stands of tubing joined by couplings 7G, each such stand being surrounded by one or more rubber centralizers 72 which are fitted snugly to the tubing, are provided with ample passages 73 for passage of the well fluid, and which engage the tubing walls to centralize the tubing therein.

Certain of the tubing couplings 7%, including usually the one nearest the lower end, are provided with check valve and fluid-impelling elements, being within a predetermined spacing range as explained hereinabove. Such a coupling is shown in typical enlarged detail in FIG. 5. The coupling collar 70 screw-threadedly joins the upset end portions on the adjacent ends of two lengths of tubing. A generally tubular valve body 74 is positioned within tubing collar 70 and within the upset end portions of the adjacent lengths of tubing. Valve body 74 is recessed on the outside to accommodate rubber ring 75 which is compressed when the coupling is made, in an obvious fashion, to position the valve body. The valve body has a central bore 78, which slidably receives a tubular stem 79 formed with a central longitudinal bore 8() to receive a long bolt S1 which clamps upper and lower rubber valve heads or disks 82 and 83, respectively, against the corresponding ends of the stem 79. Longitudinal passageways 84 are formed through valve body 74, opening into the lower end of the valve body outside rubber head 83, and opening at the top inside the rubber head S2. The conical upper surface S of the valve body through which `opens the passage 84 is, as here shown, formed on an angle which conforms to the angular lower surface of the rubber disk S2, and the latter, in moving onto and olf of surface 85, acts as a valve. Surface 85 acts as the coacting valve seat. It will be seen that the length of stem 79 is such as to permit the rubber valve head 82 to seat against valve seat 85, so as to close off passageway 84, or to be elevated sufficiently to open good flow passages from passageway 34 around the outer periphery of the valve element S2 to the tubing above.

The use of rubber having a number between the approximate range of 40 to 70 on the Shore scale for the valve element S2, and especially with a plastic such as nylon for the member affording valve seat 85, gives van impelling element which is especially effective for applying kinetic energy to the fluid stream.

The essential operation of such a sonic pump is set forth in my aforementioned patents, and need not be repeated in minute detail. Speaking generally, the oscillator G is driven by the prime mover at a frequency such as will create -a resonant standing wave in the elastic transmission line, here the elastic tubing string 12. The oscillating force applied to the generator launches alternating wave of tension and compression down the tubing, traveling in the tubing with the speed of sound, and these are reflected back up the tubing and, if the frequency of the oscillator is a resonant frequency of the tubing, the reflected wave interferes with the oncoming wave so as to produce velocity antinodes (regions of large vibratory amplitude) spaced a half-wavelength apart, and nodes (regions of minimized vibration amplitude) between the antinodes. Thus, the check valve assembly 79, `80, 82, 83 reciprocates automatically by the influence of the standing wave action in the tubing, closing the fluid passages on the up-stroke, whereby the liquid column above the unit is accelerated upwardly, and opening these passages on the down-stroke, to permit upward passage of liquid through passages 84 at this time. The movements of the valve body under the drive of the sonic |wave in the tubing occur with an acceleration many times that of gravity, and well fluids are thus pumped.

According to the special technique of the present invention the oscillator is driven by the prime mover, in this case an induction motor 36, at a frequency which is just on the low side of the frequency for peak vibration amplitude with a sonic resonant harmonic standing wave set up in -the tubing string by the oscillator. FIG. 6 shows the relationship between cyclic strain amplitude in the tubing (the vibratory stroke) and the frequency of the oscillator for a frequency range starting below and continuing past resonance. For present purposes, the hump in the curve denotes the range of resonance, and the point at which strain amplitude attains its peak maximum magnitude is denoted variably as peak resonance, peak vibration or vibratory stroke amplitude, or peak strain. In this connection, it will be appreciated -that in Iresonant standing wave vibration, each half-wavelength of the vibrating member undergoes alternating elastic elongation and contraction. The strain, then, is the amplitude of elastic elongation, or of contraction, and reaches its peak value at the peak resonance frequency, where vibration stroke amplitude attains its maximum magnitude. Accordingly, referring again to FIG. 6, the yoscillator G is driven at -a frequency to set up a resonant standing wave in the tubing, i.e. at a frequency to give performance within the region of the resonant hump of the curve as earlier described, but also, and more particularly, in a region of this hump that is below peak resonance, i.e. below the peak -of elastic strain. As represented diagrammatically in the typical diagram of FIG. 6, the periodic strain amplitude is limited to -a range such as n, well down from the peak of resonance, and yet strongly in the realm of resonance. Corresponding -to this maintained strain range of n is a frequency range r, which is held close to but safely below the resonant peak.

With the establishment vof this modified form of standing wave resonance in the pump tubing, velocity antinodes appear in the tubing, these being the maximized elastic strain regions of the tubing, but the elastic strain at these regions of the tubing is limited and below the value that would be attained at peak resonance.

lnsofar as is possible or practically feasible, the tubing couplings containing the check valve and fluid-impelling units are located with a spacing more than ve feet and less than one hundred feet, by choosing tubing lengths, as already mentioned.

It was mentioned hereinabove that a preferred feature in the practice of the invention, when gassy well fluids are pumped, is to de-gas the well fluids, either prior to entry of the ywell fluids into the pump tubing, or during transit up the pump tubing. By installing the pump so that its open lower end is well below the casing perforavtions through which production fluids enter the well, e.g.,

those located at 13a in FIG. 1, there can be a considerable liquid column such as indicated at 9), for example, and such a column will give off a large portion of its gas, to rise in the well annulus, and t0 be withdrawn via annulus port 50 and a flow line (not shown) connected to the latter. A high liquid column can also be maintained in the Well annulus even without the relatively high perforations. Also, liquid may be introduced to the annulus from above, through the annulus port 5f?, to provide a high liquid column. Such high liquid column, by the pressure head developed, causes the gas to rise and go olf the top. Thus the well fluids are subjected to dei gassing prior to entry into the -tubing inlet nipple 12a forming the lowermost segment of the pump tubing.

Also, by the step of installing the pump tubing quite deep relative to the producing formation and the perforations, there is a downow of the well fluids around the outside of the vibrating tubing prior to entry of the well `fluids into the tubing string. By ythe sonic activity of the tubing in this outside contact with the well uids, the well uids become further de-gassed before entry into the tubing. Finally, gas bleed holes 94 in the tubing de-gas the liquid as `it goes up the tubing, after entry.

To `attain resonant harmonic standing Wave vibration in the tubing string and a vibration frequency just below that lfor the peak of resonance, the electric drive motor 36, which may be an induction motor, is karranged to drive the oscillator and tubing string at that frequency, i.e. just below the resonance peak, where the elastic stress and strain in the tubing are just under the resonantly amplified maximum that would occur at peak resonance. An induction motor has a sufficient degree of inverse speed responsiveness to load to behave as desired in the system. In -carrying out the process of the present invention, the sonic wave in the tubing string is, -rst of all, arranged to be near that of a harmonic of the fundamental resonant frequency of the tubing. This clearly requires that there be used an induction motor, or other prime mover, which will drive the oscillator and the vibratory tubing string in the frequency range of a resonance hump, as explained above. A refined frequency adjustment is then made to establish vibration lfrequency near to peak resonance, but on the low side thereof. Thus, in other language, the frequency of the periodic elastic strains created in the tubing is near to but under the frequency for maximum resonant amplitude of said strains at the harmonic at which operation is taking place. Still further, adjustments are made so as to establish the amplitude of the elastic strains to -a magnitude corresponding to a stress amplitude which does not exceed that lfor the endurance limit of the tubing in the environment of the well at the operating frequency. It will be seen that the process is carried out by using a prime mover, with a speed rating and a drive ratio land power output to the oscillator, such that, whatever the load on the prime mover involved in vibrating the tubing string and elevating the column of Well duid, the prime mover drives the oscillator in the desired harmonic frequency range, on the low side of the resonance peak. Then, if load on the oscillator falls off, as by gas in the oil column in the tubing, the motor, owing to its inverse speed responsiveness to load, tends to increase its speed, thereby increasing the frequency of the oscillator, and the frequency of the standing wave in the elastic column, i.e., in this case, the pump tubing. Thereby, the well liquid is pumped up the tubing at faster rate, and the oscillator regains loading. 'This increase in loading takes place sharply and rapidly, and the operation stabilizes out at a steady frequency again on the low side of the resonance peak.

Also, the unbalanced weights of the mass-reactive oscillator G, and their centers of gravity, are so adjusted relative to the cross-sectional area of the column or tubing that, with the angular velocity of these weights limited to the value at which, with the oscillator operating at the predetermined resonance tfrequently just under the peak of the selected resonance hump, the stress amplitude is limited to a value 0f 20,000 p.s.i. In this connection, the 20,000 p.s.i. limit here referred to is the permitted dep-arture, compressive or tensional, from neutral, and is the half-amplitude of the total permitted cyclic stress swing.

The method of the invention also, to safeguard the pump tubing or other sonic transmission column from elastic fatigue failure, involves an adjustment of the various inter-related factors involved, such as the alternating output force from the oscillator, and the crosssectional area of the tubing or other column, such that the stress and strain amplitudes at the below-peakresonance frequency of operation do not exceed those for the endurance limit of the tubing in the environment of the well at said frequency.

Certain illustrative embodiments of the invention have now been described and illustrated, but it will be understood that these are for illustrative purposes only, and that various changes in design, structure and equipment may be made Without departing from the spirit and scope of the appended claims.

I claim:

1. In the operation of a sonic deep well pump having a well fluid conduit with iluid-impelling means therein reciprocated by periodic longitudinal elastic strains in a distributed-constant sonic wave transmission line by propagation of a sonic wave along said line, with said line coupled to said iluid-impelling means, the method that comprises:

establishing sonic wave propagation along said transmission line at a frequency which is near a harmonic of the fundamental resonant Afrequency of said line, so as to cause said periodic elastic strains in said line at a resonantly magnified amplitude;

adjusting the frequency of said elastic strains to a frequency which is near to but under the frequency for `maximum resonant 'amplitude of said strains at said harmonic; and

adjusting the amplitude of said elastic strains to a magnitude which is no greater than that for the endurance limit of said elastic transmission line in the environment of said well at said adjusted frequency.

2. The method of claim 1, wherein said strain amplitude is no greater than that produced by a stress amplitude of twenty thousand pounds per square inch in said elastic transmission line.

3. The method of claim 1, including the step of degassing the pumped liquid.

4. The method of claim 3, wherein said de-gassing process is accomplished vbefore the liquid enters said conduit.

5. The method of claim 3, wherein said de-gassing process is applied to said liquid 'as it is impelled along a portion of said conduit.

6. The method of claim 1, including locating a plurality of uid-impelling means along a portion of said conduit with an average spacing of more than ve feet and less than one hundred feet.

7. The method of claim 1, wherein said adjusted frequency is no less than four cycles per second, and no greater than one hundred cycles per second.

8. The method of claim 1, wherein said elastic transmission line is Operated by driving it sonically from a mass-reactive oscillator employing rotating unbalanced weights; and

rotating said weights by lturning a drive shaft directly coupled to said weights, and driven by a prime mover.

9. The method of claim 8, including also stabilizing said drive shaft by means of a flywheel.

References Cited by the Examiner UNITED STATES PATENTS LAURENCE V. EFNER, Primary Examiner, 

1. IN THE OPERATION OF A SONIC DEEP WELL PUMP HAVING A WELL FLUID CONDUIT WITH FLUID-IMPELLING MEANS THEREIN RECIPROCATED BY PERIODIC LONGITUDINAL ELASTIC STRAINS IN A DISTRIBUTED-CONSTANT SONIC WAVE TRANSMISSION LINE BY PROPAGATION OF A SONIC WAVE ALONG SAID LINE, WITH SAID LINE COUPLED TO SAID FLUID-IMPELLING MEANS, THE METHOD THAT COMPRISES: ESTABLISHING SONIC WAVE PROPAGATION ALONG SAID TRANSMISSION LINE AT A FREQUENCY WHICH IS NEAR A HARMONIC OF THE FUNDAMENTAL RESONANT FREQUENCY OF SAID LINE, SO AS TO CAUSE SAID PERIODIC ELASTIC STRAINS IN SAID LINE AT A RESONANTLY MAGNIFIED AMPLITUDE; ADJUSTING THE FREQUENCY OF SAID ELASTIC STRAINS TO A FREQUENCY WHICH IS NEAR TO BUT UNDER THE FREQUENCY FOR MAXIMUM RESONANT AMPLITUDE OF SAID STRAINS AT SAID HARMONIC; AND ADJUSTING THE AMPLITUDE OF SAID ELASTIC STRAINS TO A MAGNITUDE WHICH IS NO GREATER THAN THAT FOR THE ENDURANCE LIMIT OF SAID ELASTIC TRANSMISSION LINE IN THE ENVIRONMENT OF SAID WELL AT SAID ADJUSTED FREQUENCY. 